The present invention relates to vehicle transmissions, especially for heavy, on- and off-road vehicles, and more particularly to dual- and multi-clutch transmissions with a range section that, as set out in the preamble of claim 1 and as disclosed in U.S. Pat. No. 5,511,437.
Dual clutch transmissions are a cross-breed between conventional stepped transmissions, with power interruption at gear shifts, and powershifting, without power interruption, planetary transmissions. In principle, a dual clutch transmission has two input shafts, each connectable to a friction clutch and to the output of the engine. Functionally, this is equivalent to having two conventional transmissions in parallel, i.e., two parallel sub-transmissions, and using one at a time for poller transfer. The sub- transmission that is not used, idling, for the time being! can have a gear engaged and prepared for a subsequent shift. This shift is carried out by simultaneously disengaging the friction clutch of the previously used sub-transmission and engaging the friction clutch of the previously idling sub-transmission.
When properly designed, dual clutch transmissions have a potential of providing powershifts at a reasonable production cost and low power losses. This is due to the fact that the rotating parts, i.e., gearwheels, shafts and tooth clutches, are similar to those in conventional stepped transmissions. This, furthermore, enables the use of the same production equipment. So, it makes sense to produce dual clutch transmissions in the same facilities as used for conventional stepped transmissions.
Dual clutch transmissions for rear wheel drive vehicles often have two separate countershafts, one connected to each input shaft. One example is found in U.S. Pat. No. 5,15,0628 (referred to as 12 and 15). These countershafts are parallel to the transmission input. They make the transmission wider than a conventional stepped transmission. That may lead to difficulties in installing the transmission into the vehicle. On the other hand, the transmission may be shorter instead. Anyway, in some dual clutch transmission designs there is only one countershaft, e.g., as in DE923402 and DE3131156A1. On this countershaft there are loose gearwheels arranged that can be rotationally connected to each other and to the countershaft by selective engagement of mechanical tooth clutches. In a way, this could be regarded as if the second countershaft is arranged coaxial to the first one. The result is a powershiftable dual clutch transmission that is as slim as corresponding conventional stepped transmissions. The number of gears and the speed reduction ratios possible are insufficient for heavy duty vehicles, though.
Some dual-clutch transmission concepts in a so-called winding structure have been presented, for instance U.S. Pat. Nos. 5,347,879, 5,592,854, DE10325647A1 and DE10339758A1. In these, the power is led via four gear meshes in at least one gear, and several gearwheels are used for more than one gear. That will give further reduction of speed. However, this corresponds to just one or two additional gears. These concepts are, hence, less suited to heavy vehicles.
DE102005030987A1, DE102005033027A1 , DE102006015661 A1 and EP1624232A1 show transmission concepts where a main transmission of dual clutch type is connected in series with a range section. That makes it possible to double the number of gears and obtain gears for high tractive force as well as gears tor high vehicle speed.
Unfortunately, there are shifts between consecutive gears where the power transfer will be interrupted in these designs. That is not an option for heavy on- and off road vehicles subjected to high driving resistance. Two similar designs of dual clutch main transmission in combination with a range section are shown in
DE102005050067A1 and WO2007/039021A1. Therein, the input and output of the transmission can be connected by a friction clutch. This friction clutch can transfer power when a gear shift takes place. Thereby, power interruption can be avoided at all gear shifts.
However, for reasonable sizes of this friction clutch, the power transferred to the driven wheels is very small at shifts between low gears. At the same time, the power dissipated in this friction clutch is large at these gear shifts. Thus, these types of dual clutch transmission would have a limited practical use, especially for on- and off-road vehicles.
Dual-clutch transmissions as in U.S. Pat. No. 5,150,628, DE923402 and DE3131156A1 could be combined with a range section.
That gives a compact transmission with several gears and high speed reduction ratios. Gear-shifts between consecutive gears could be without power interruption, except when, the range section is shifted. This would probably be acceptable on most heavy on-road vehicles, but not for, e.g., trucks in hilly applications or articulated haulers.
U.S. Pat. No. 7,353,724 B2 shows in FIGS. 1 and 3 dual-clutch transmissions where a direct connection between one of the input shafts and the output shaft can transmit power when changing between low, underdrive, gears and high, overdrive, gears. This is not a true range section, though. The number of gears is doubled, but in the underdrive gears the power is transmitted via two gear meshes, only, as in FIG. 2 in U.S. Pat. No. 6,958,028 R2. That limits the practically possible speed reduction.
Thereby, these transmissions are not suited to heavy vehicles.
A somewhat similar principle is disclosed in U.S. Pat. No. 4,777,837. There, separate gearwheel pairs are provided for intermediate gears between the low and high range gears. This will give a large number of gears and no power interruptions at gear-shifts between consecutive gears. In low range gears, the power is transmitted via three gear meshes, which will enable large reduction ratios. However, the transmission is both wide and long due to two parallel countershafts and a large number of gearwheels located, side by side. Moreover, the output shaft is not coaxial with the input shaft. That makes the transmission incompatible with most heavy truck designs. The number of components is large, adding costs.
Further on, U.S. Pat. No. 7,070,534 B2 presents a dual clutch transmission 10 with a planetary range section 56 and coaxial input 86 and output 68. A dual clutch unit 20,22 selectively transfers power to input shafts 90 and 92. To each of these input shafts 90, 92 a countershaft, 74 and 76, respectively, is arranged. From each countershaft 74, 76 the power can be selectively directed with tooth clutches 80 and 84 to the output 68 in either of two ways. Firstly, the power can be led to the sun gear 58 of the planetary range section 56 via gearwheels 44, 46 and 54, 46, respectively. That will give a speed reduction in the planetary range section 56, corresponding to low range gears. Secondly, the power can be led more directly to the output 68 via gearwheels 40, 42 and 50, 42, respectively. The planetary range section will then be idling, and high range gears are established. Shifts without power interruption can be carried out between low and high range gears. Unfortunately, the number of rotating components, e.g., gearwheels and tooth clutches, is relatively large in comparison with the number of gears. The large number of gearwheels makes the transmission long, and the two parallel countershafts make it wide. Thereby, it will be difficult to fit in the vehicle. The transmission will be costly to manufacture due to the large number of components. Furthermore, the idling planetary range section will imply unnecessarily large power losses in high range gears. Hence, there are several disadvantages that make this transmission less suited for use in heavy vehicles.
U.S. Pat. No. 6,958,028 B2, FIG. 5, shows a dual clutch transmission with a planetary range section. This transmission is similar to the one in U.S. Pat. No. 7,070,534 B2. The main difference is that both input shafts, 30 and 40, use the same countershaft 50, tooth clutch 130, and gearwheels 122, 124 and 126, 128 between this countershaft and the planetary range section. Power interruption between low and high range gears is eliminated by a bridge torque path via a separate countershaft 152. That makes the transmission wide, and it shares the rest of the disadvantages of the one in U.S. Pat. No. 7,070,534 B2; many components, long, and high power losses for high range gears.
In US 2008/0188342A1, FIG. 1 presents a single countershaft dual clutch main transmission 10 combined with a planetary range section 12. A bridge torque path is formed by a tooth clutch 84 that rotationally locks a loose gearwheel 64 on main shaft 28 to a planet carrier 68 rotationally fixed to output shaft 70. When power is led in this path, the gearwheels in the planetary range section are idling, and it can be shifted between high and low positions. This gives a narrow transmission with high reduction ratios where power interruptions can be avoided at every shift between consecutive gears. However, this bridge path embodiment has a number of drawbacks. Firstly, the tooth clutch 84 is of complex design, making it costly and long. Secondly, the bearing 32 that carries main shaft 28 must be located in front of loose gearwheel 64. This puts a large part of the main shaft 28 behind bearing 32, which, in turn, increases misalignments in the ramie section and tooth clutch 84. Moreover, the assembly of the transmission is not facilitated by a main shaft having gearwheels and tooth clutches on both, sides of the housing wall that carries bearing 32.
Thirdly, adding parts for tooth clutch 84 will make the already complex shape of planet carrier 68 even more complex and tricky to produce. FIG. 3 shows a similar dual, clutch main transmission in combination with a non-planetary range section 102. For the rest, this transmission has similar properties as the one in FIG. 1. DE102007047671A1 shows a similar design that has similar disadvantages.
U.S. Pat. No. 5,385,066 presents several transmissions where a conventional stepped main transmission 12 is combined with a non-planetary range section 14. Tooth clutches 184 can selectively rotationally lock the main transmission countershafts 142 to the range section countershafts 172. Thereby, additional gears can be obtained without adding gearwheels. That will make the transmissions short in length and cost-effective to produce. However, these transmissions are of conventional type, having a single frictional master clutch and power interruption at every gear shift. This limits their use to on-road, non-severe heavy vehicle applications.
Thus, for heavy, on- and off-road vehicles there is a need for a transmission that i) enables high power transfer to the driven wheels during all shifts between consecutive gears, it) can provide high reduction ratios, iii) is cost-effective and simple to produce, iv) has low power losses and v) can be installed in a vehicle where the space available, especially axially, is limited.
Thus, it is desirable to present an improved transmission.